1. Field of the Invention
The present invention relates to a hydraulic device used for a hydraulic excavator for a construction machine or the like. The hydraulic device is adapted for controlling the delivery oil from one or more hydraulic pump(s) which flows into and drives both at least one actuator having an excessively higher inertial load and at least one actuator having a relatively low inertial load at the same time.
2. Description of the Related Art
This type of hydraulic device is employed primarily for construction machinery and agricultural machinery. It is equipped with a load-sensing required-stream regulation function for controlling the delivery of the variable displacement pump according to loaded pressure. Further, the circuits connected to actuators are provided with pressure compensation valves to divide the pump delivery so as to prevent the respective actuators from interfering with each other due to the difference in loaded pressures, etc. among the respective actuators with a resultant change in speed of the actuators when driving the plurality of actuators at the same time. Furthermore the hydraulic devices are equipped with a function known as an anti-saturation function for distributing pump delivery to the individual actuators at an appropriate ratio when the pump delivery is smaller than a predetermined required flow of the plurality of driven actuators.
A first such conventional hydraulic device is shown in FIG. 5 which is disclosed, for instance, in U.S. Pat. No. 5,347,811, Japanese Publication No. 05172112; 08254201. In FIG. 5, a so-called `after-orifice type` hydraulic device having a load-sensing function is shown which comprises first and second actuators 12,13 and first and second directional valves 14,15 each having flow control function capable of controlling the pump delivery oil from a variable delivery pump 2 flowing into each of the actuators, respectively, and first and second pressure compensation valves 50,51 coupled to and for compensating pressures of the first and second directional valves 14, 15, respectively. Each pressure compensation valve 50,51 located between the actuator and the directional valve both communicating with the pressure compensation valve receives an oil pressure on a downstream side of a throttle of the directional valve coupled to the pressure compensation valve to act in a first control pressure chamber 50a,51a to open the pressure compensation valve and a maximum loaded pressure of the loaded pressures of the hydraulic actuators of the hydraulic device to act in a second control pressure chamber 50b, 51b to close the pressure compensation valve, and a pressure receiving area of each control pressure chamber 50a,51a,50b,51b is made nearly equal with each other.
With this arrangement, on condition that the amount of the delivery oil flow to be supplied to each actuator 12,13 is relatively small and the total amount to be supplied to each actuators does not reach to the maximum delivery flow rate of the pump 2, a differential pressure across each directional valve 14,15 across its throttle becomes equal to the differential pressure between the pump delivery pressure Pp and the maximum loaded pressure Pm among the actuators, that is, equal to the pre-set differential pressure being set by the spring 18 of the pump flow control valve 17. Therefore, even if the load pressures of the actuators 12,13 may differ from each other, each differential pressure across each directional valve will not be affected by the load pressure of each actuator, thereby the amount of the delivery oil flow to be supplied to each actuator 12,13 is determined by an amounts of the openings of the throttles of the directional valves and the pre-set differential pressure being set by the spring 18, and performs a load-sensing function to keep a pre-set speed control of the actuators. Further, the maximum pressure Pm of the actuators is introduced to the pump flow control valve 17 to drive the displacement varying means 6 coupled to the pump 2, so that the differential pressure between the pump delivery pressure Pp and the maximum loaded pressure Pm is controlled to be equal to the pre-set differential pressure being set by the spring 18.
In FIG. 5, assuming that the first actuator 12 being for a swing motor for a cab for a hydraulic excavator of a construction machine having a high inertial load, and the second actuator 13 being for a boom cylinder having a low-load, and these actuators are simultaneously operated. Firstly, control levers of the directional valves 12,13 are moved by certain strokes to operate the actuators. These strokes are usually of long and full or nearly full ones. Then, the delivery oil from the pump 2 flows into the actuators 12,13 through the directional valve 14,15 to move the actuators. However, since the actuator 12 has the high inertia, the actuator 12 does not immediately move, thus causing a loaded pressure of one of the inlet ports 12a,12b to rise momentarily. An excessive rise in the loaded pressure of the one of the inlet ports exceeding a relief-setting pressure of the overload relief valves (not shown) connected to the lines of the inlet ports 12a,12b, further causes a rise of the loaded pressure of the one of the relief valve, thereby almost of the delivery oil flowing into the one of the inlet port is exhausted through the relief valve into a tank T.
At the same time, since the loaded pressure exceeding over the relief-setting pressure is introduced, by way of the load-sensing function, to the pump flow control valve 17 through a shuttle valve 4 via line 5 to act the pump flow control valve 17 to increase pump delivery oil. On the other hand, when the delivery oil pressure of the pump 2 rises to a predetermined setting pressure of a constant power output regulation valve 19, the valve 19 take precedence over the pump flow control valve 17 to decrease the delivery oil from the pump 2 through the auto-constant power output regulation function.
While the constant power output regulation valve 19 is acting to decrease the delivery oil from the pump 2, the above mentioned `anti-saturation function` for distributing pump delivery to the individual actuators at an appropriate ratio when the pump delivery is smaller than a predetermined required flow of the plurality of driven actuators, works to act on the pressure compensation valve to keep each oil pressures on the upstream side lines 7 and 7 of the pressure compensation valves 50,51 to be equal. This results that each opening of the throttles of each pressure compensation valve is made smaller and the delivery oil flows through the pressure compensation valves will decrease. Therefore, the speed of the boom cylinder 13 becomes extremely slower than that of when the boom cylinder 13 is independently operated, causing the boom cylinder operation such as a loading on a truck excessively difficult, deteriorates the working efficiencies and increases the operator's fatigue. At the same time, a problem occurs that the delivery oil flow flowing into the actuator 12 for the swing motor and then exhausted through the overload relief valve into the tank causes a large energy loss of the engine 1.
Secondly, when actuator 12 for the swing motor loses the acceleration and reaches to a constant speed operation, the torque of the swing motor suddenly decreases, then the loaded pressure of the actuator 13 for boom cylinder becomes higher than that of the actuator 12 by the lowering of its loaded pressure. This causes the maximum loaded pressure Pm of the hydraulic actuators on the line 5 suddenly drops and thereby results a drop of the line pressure on the pump delivery line 3 and increases the pump delivery oil through the easing of the operation of the constant power output regulation valve 19, resulting that the speed of the actuator 13 for boom cylinder is suddenly accelerated, and as a whole, the actuators 12,13 do not work smoothly during the simultaneous operations of these actuators 12,13.
To cope with these problems of the first conventional hydraulic device shown in FIG. 5, U.S. Pat. No. 5,347,811 and the Japanese Publication No. 08254201, for example, propose a hydraulic device wherein a downstream line of a pressure compensation valve for an actuator for a swing motor and an inlet port of an actuator for extending the actuator for the boom cylinder is communicated with each other via a joinning line , and th ere a re pro vided on the joining line in series a pilot operate shut-off valve and a check valve allowing a flow to the inlet port from a downstream line of the pressure compensation valve for the swing motor. In operation, in accordance with the amount of the strokes of the directional valves, a pilot pressure of the directional valve for the boom cylinder is introduced to open the shut-off valve and the check valve in series, thereby the maximum loaded pressure on a downstream line of the pressure compensation valve for the swing motor flows into the inlet port of the actuator for extending the boom cylinder. This prevents an abrupt rise of the load pressure of the actuator for the swing motor and at the same time prevents the lowering of the extending speed of the boom cylinder.
However, to prevent both the abrupt rise of the load pressure of the actuator for the swing motor and the lowering of the extension speed of the boom cylinder, the afore-mentioned U.S. Pat. No. 5,347,811 and the Publication No. 08254201 must provide beside the conventional pressure compensation valves the additional valves, such as the pilot operate shut-off valve, the check valve, and the external pilot lines providing pilot pressures to operate these additional valves at a predetermined condition. Therefore, the additional valves and the external pilot lines naturally make total valve block and hydraulic system bulky, complicated and of high cost. Further, since these additional valves operates at the predetermined condition, an additional problem occurs that the boom cylinder makes a discontinuous movement.
The above-mentioned problems of the first conventional hydraulic device shown in FIG. 5 similarly occur in a second conventional hydraulic device shown in FIG. 6 which is disclosed, for example, in the Japanese Publication No. 07324355. In FIG. 6, a hydraulic circuit for a hydraulic device having both the anti-saturation function and the load-sensing function is shown comprising first and second directional valves 24,25 disposed in parallel each having flow control function capable of controlling the pump delivery oil from a variable displacement pump 2 flowing into each of actuators 12,13 via a pump line 3 and check valves 26,27, respectively. First and second pressure compensation valves 60,61 for compensating pressures of the first and second directional valves 24,25 are located on downstream sides of the directional valves 24,25 before a tank T, respectively. Each return oil flowing from the actuators 12,13 is exhausted via the directional valve 24, 25, the pressure compensation valve 60, 61 and tank line 16 to the tank T. Each pressure compensation valve 60,61 receives an oil pressure communicated with a loaded pressure of the actuator communicating with the pressure compensation valve to act in a first control pressure chamber 60a,61a to open the pressure compensation valve, and a maximum loaded pressure of the loaded pressures of the hydraulic actuators of the hydraulic device to act in a second control pressure chamber 60b,61b to close the pressure compensation valve, respectively. A pressure receiving area of each control pressure chamber 60a,61a,60b,61b is made nearly equal. By such an arrangement, this second conventional hydraulic device shown in FIG. 6 performs similar operations and has the same problems as described in the first conventional hydraulic device shown in FIG. 5.
To cope with these problems of the second conventional hydraulic device shown in FIG. 6, the Japanese Publication No. 07324355, for example, proposes a hydraulic device wherein, adding to the hydraulic circuit shown in FIG. 6, a bypass pump delivery oil line communicating with the tank is provided in parallel to the directional valves, and a bleed-off valve and a pressure generating device are provided in the bypass pump delivery oil line in series. And a pressure on an upstream side of the pressure generating device is introduced to the pump displacement varying means coupled to the variable displacement pump to perfrom a so-called a negative control. Further, a maximum pressure of all actuators is adapted to act only on the pressure compensation valve coupled to the actuator for a swing motor and on the bleed-off valve to close the pressure compensation valve and the bleed-off valve, while a maximum pressure of actuators other than for a swing motor having a relatively low load is adapted to act only on the pressure compensation valves coupled to the actuators other than for the swing motor having a relatively low load to close the pressure compensation valves, thereby the pressure compensation valves coupled to the actuators other than for the swing motor are prevented from closing the pressure compensation valves by an excessive high loaded pressure of the swing motor and are prevented from decreasing the moving speed of the actuators other than for the swing motor having a relatively low load. However, the above addition of the additional valve and the pilot lines to operate the pressure compensation valves for the boom cylinders naturally make the hydraulic circuit complicated and total valve blocks bulky, and of high cost. Furthermore, when actuator for the swing motor loses the acceleration and reaches to a constant speed operation, the load of the swing motor suddenly decreases, then the loaded pressure of the actuator for boom cylinder becomes higher which causes the pressure compensation valves for the actuator for the swing motor to close by the high loaded pressure of the actuator for boom cylinder, resulting the sudden lowering of the swing speed of the actuator for the swing motor.
The above-mentioned problems of the first conventional hydraulic device shown in FIG. 5 similarly occur in a third conventional hydraulic device shown in FIGS. 7 and 8 which is disclosed, for example, in U.S. Pat. No. 5,622,206, and Japanese Publication No. 05332310; 05332311. In FIG. 7, a hydraulic device is shown which comprises pressure compensation valves 70,71 located between pump lines 3 and directional valves 24,25 communicating with the pressure compensation valves 70, 71, respectively. Each pressure compensation valve 70, 71 is integrally formed with a check valve portion 76, 78 which normally blocks the reverse flow from the actuator to the pump lines 3 and throttles the pump delivery oil flowing into the actuator, and a reducing valve portion 77,79 having a reducing valve spool 72 contactable to close the check valve spool 74 of the check valve portion 76,78 and capable of reducing a pressure of the pump delivery oil on the pump lines 3 to a loaded pressure of the actuator communicating with the reducing valve portion 77,79, respectively. And each pressure compensation valve 70,71 receives an oil pressure on a downstream side of a throttle of the directional valve 24,25 coupled to the pressure compensation valve to act in a first control pressure chamber 77a, 79a of the pressure compensation valve to open the pressure compensation valve, a maximum loaded pressure of the loaded pressures of the hydraulic actuators of the hydraulic device to act in a second control pressure chamber 77b,79b of the pressure compensation valve to close the pressure compensation valve, respectively. A pressure receiving area of each control pressure chamber 77a,79a,77b,79b is made nearly equal. FIG. 8 is a schematically cross sectional block view of one of the pressure compensation valves 70,71 shown in FIG. 7. By such an arrangement, this third conventional hydraulic device shown in FIG. 7 performs similar operations and has the same problems as described in the first conventional hydraulic device shown in FIG. 5.
To cope with these problems of the third conventional hydraulic device shown in FIG. 7, to prevent the lowering of the extension speed of the boom cylinder having a relatively low inertial load, the Japanese Publication No. 05332311, for example, proposes a hydraulic device wherein, beside the conventional pressure compensation valves and the directional valves, a pilot check valve is provided which opens by an introduction of a pilot pressure on a pilot pressure line adapted to move the spool of a directional valve 25 communicated with an actuator 13 for a boom cylinder having the relatively low inertial load. The pilot check valve is located before the reducing valve portion 77 communicated with the actuator 12 for the swing motor, and thus prevents an introduction of the pump delivery oil to the reducing valve portion 77 communicated with the actuator 12 for the swing motor, thereby prevents both the abrupt rise of the load pressure of the actuator for the swing motor and the lowering of the extension speed of the boom cylinder.
However, the Japanese Publication No. 05332311 must provide beside the conventional pressure compensation valves the additional valve such as the pilot operate check valve, and the pilot line to operate the pilot operate check valve. Therefore, the additional valves and the pilot line naturally make the hydraulic circuit complicated and total valve blocks are bulky, and of high cost.